1. Field of the Invention
A continuously variable transmission apparatus according to the present invention is used as a transmission unit constituting an automatic transmission for a vehicle or as a transmission for any one of various kinds of industrial mechanical apparatus. Especially, a continuously variable transmission apparatus according to the present invention is developed in order to reduce an incongruous feeling given to a driver when switching a low speed mode and a high speed mode over to each other using a clutch mechanism.
2. Description of the Related Art
As an automatic transmission for a vehicle, such a toroidal-type continuously variable transmission as schematically shown in FIGS. 10 and 11 is known well because it is disclosed in many publications such as patent publications, and it is also enforced currently in part of the vehicle industry. When a toroidal-type continuously variable transmission of this type is in operation, as an input shaft 1 is rotated, a pressing device 2 such as a loading cam rotates an input side disk 3 while pressing the input side disk 3 against a plurality of power rollers 4, 4. And, the rotational power of the input side disk 3 is transmitted through the plurality of power rollers 4, 4 to an output side disk 5, so that an output shaft 6 fixed to the output side disk 5 is rotated.
To change the rotation speed between the input shaft 1 and output shaft 6, trunnions 7, 7 supporting the power rollers 4, 4 thereon are respectively swung about their associated pivot shafts 8, 8 which are disposed on the two end portions of each of the trunnions 7, 7 in such a manner that the pivot shafts 8, 8 are respectively concentric with their associated trunnions 7, 7. In this case, in case where the rotation center axes of the power rollers 4, 4 are swung in such a manner that the peripheral surfaces 4a, 4a of the power, rollers 4, 4, as shown in FIG. 10, can be respectively contacted with the near-to-center portion of the inner surface 3a of the input side disk 3 and the near-to-outer-periphery portion of the inner surface 5a of the output side disk 5, the rotation speed between the input shaft 1 and output shaft 6 can be reduced. On the other hand, in case where the rotation center axes of the power rollers 4, 4 are swung in such a manner that the peripheral surfaces 4a, 4a of the power rollers 4, 4, as shown in FIG. 11, can be respectively contacted with the near-to-outer-periphery portion of the inner surface 3a of the input side disk 3 and the near-to-center portion of the inner surface 5a of the output side disk 5, the rotation speed between the input shaft 1 and output shaft 6 can be increased.
While the basic structure of a toroidal-type continuously variable transmission is as described above, to swing the trunnions 7, 7 about their respective pivot shafts 8, 8 in order to change the transmission ratio, the trunnions 7, 7 may be shifted slightly in the axial direction (in FIGS. 10 and 11, in the front and rear direction of the sheets of FIGS. 10 and 11) of the pivot shafts 8, 8. As the trunnions 7, 7 are shifted in this manner, the direction of a force to be applied to the rolling contact portions between the above-mentioned surfaces 3a, 4a and 5a in the tangential direction thereof is caused to vary, with the result that trunnions 7, 7 can be swung. The structure of this portion is also well known because it is disclosed in many publications such as patent publications, and the present structure is currently enforced in part of the vehicle industry. Also, structures, each of which is used to increase the number of power rollers for use in power transmission to thereby be able to increase the power that can be transmitted by the toroidal-type continuously variable transmission, are also disclosed in many publications such as patent publications and thus are known well; and, part of these structures are now practically applied. As such structures for increasing the number of power rollers in this manner, for example, there are widely known a structure in which two pairs of input side and output side disks are disposed in parallel to each other with respect to the transmission direction of the power, and a structure in which the number of power rollers to be interposed between the input side and output side disks can be increased.
Further, as disclosed in JP-A-1-169169, JP-A-1-312266, JP-A-10-196759, JP-A-11-63146, JP-A-11-63148, and JP-A-2000-220719, conventionally, there is also proposed a structure in which a toroidal-type continuously variable transmission and a planetary gear transmission mechanism are combined together to thereby construct a continuously variable transmission apparatus. The present continuously variable transmission apparatus is invented so as to increase the width of the transmission ratio; or, to reduce the torque that passes through the toroidal-type continuously variable transmission when a vehicle is running at a high speed, thereby being able to enhance the durability of the toroidal-type continuously variable transmission; or, to allow a continuously variable transmission by itself to stop the output shaft even during the rotation of the input shaft, thereby being able to eliminate the need for provision of a start clutch.
Now, FIG. 12 shows an example of a conventional continuously variable transmission apparatus of the above-mentioned type which is disclosed in the above-cited JP-A-11-63146. This continuously variable transmission apparatus comprises a double cavity type of toroidal-type continuously variable transmission 9 and a planetary gear mechanism 10; and, specifically, in the toroidal-type continuously variable transmission 9, two pairs of input side and output side disks are disposed in parallel to each other with respect to the power transmission direction. And, according to this continuously variable transmission apparatus, in the low speed running operation, the power is transmitted only by the toroidal-type continuously variable transmission 9; and, in the high speed running operation, the power is transmitted mainly by the planetary gear mechanism 10 and, at the same time, the transmission ratio given by the planetary gear mechanism 10 can be adjusted by changing the transmission ratio of the toroidal-type continuously variable transmission 9.
For the above purpose, the base end portion (in FIG. 12, the right end portion) of an input shaft 1a passing through a centrosphere of the toroidal-type continuously variable transmission 9 and supporting a pair of input side disks 3, 3 on the two end portions thereof is connected to a transmission shaft 13, which is fixed to the central portion of a support plate 12 supporting a ring gear 11 constituting the planetary gear mechanism 10, by a high speed clutch 14. By the way, of the pair of input side disks 3, 3, the input side disk 3 disposed on the leading end side (in FIG. 12, the right side) of the input shaft 1a is supported on the input shaft 1a in such a manner that it is rotatable in synchronization with the input shaft 1a as well as it is prevented against substantial movement in the axial direction of the input shaft 1a. On the other hand, the input disk 3 on the base end side (in FIG. 12, the left side) is supported on the input shaft 1a in such a manner that it can be rotated in synchronization with the input shaft 1a as well as can be moved in the axial direction of the input shaft 1a. 
Also, between the output side end portion (in FIG. 12, the right end portion) of a crankshaft 16 of an engine 15 serving as a drive source and the input side end portion (that is, the base end portion; in FIG. 12, the left end portion) of the input shaft 1a, there are interposed a start clutch 17 such as a torque converter or an electromagnetic clutch and an oil pressure type of pressing device 18 in such a manner that they are connected in series with each other with respect to the power transmission direction. This pressing device 18 is structured by inserting the base end portion of the input disk 3 into a cylinder 19 in such a manner that the input disk 3 is oiltight and is able to transmit the rotational power thereof.
Also, an output shaft 20, which is used to take out the power based on the rotational motion of the input shaft 1a, is disposed so as to be concentric with the input shaft 1a. And, on the periphery of the output shaft 20, there is disposed the planetary gear mechanism 10. A sun gear 21, which constitutes the planetary gear mechanism 10, is fixed to the input side end portion (in FIG. 12, the left end portion) of the output shaft 20. Therefore, the output shaft 20 can be rotated as the sun gear 21 is rotated. On the periphery of the sun gear 21, there is supported the ring gear 11 in such a manner that it is concentric with the sun gear 21 and can be rotated. And, between the inner peripheral surface of the ring gear 11 and the outer peripheral surface of the sun gear 21, there are interposed a plurality of planetary gear sets 23, 23 each of which comprises a pair of planetary gears 22a and 22b combined together. Each pair of planetary gears 22a and 22b are meshingly engaged with each other; and, the planetary gear 22a disposed on the outside diameter side is meshingly engaged with the ring gear 11, while the planetary gear 22b disposed on the inside diameter side is meshingly engaged with the sun gear 21. The thus arranged planetary gear sets 23, 23 are rotatably supported on one side surface (in FIG. 12, the left side surface) of a carrier 24. Also, the carrier 24 is rotatably supported on the middle portion of the output shaft 20.
And, the carrier 24 is connected to a pair of output side disks 5, 5 constituting the toroidal-type continuously variable transmission 9 by a first power transmission mechanism 25 in such a manner that the rotational power can be transmitted between them. The first power transmission mechanism 25 comprises a transmission shaft 26 disposed so as to extend in parallel to the input shaft 1a and output shaft 20, a sprocket 27a fixed to one end portion (in FIG. 12, the left end portion) of the transmission shaft 26, a sprocket 27b connected to the respective output disks 5, 5, a chain 28 interposed across the two sprockets 27a and 27b, and first and second gears 29, 30 which are respectively fixed to the other end (in FIG. 12, the right end) of the transmission shaft 26 and the carrier 24 and are meshingly engaged with each other. Thanks to this, as the respective output side disks 5, 5 are rotated, the carrier 24 can be rotated in the opposite direction to the output side disks 5, 5 at the speed that corresponds not only to the number of teeth of the first and second gears 29, 30 but also to the number of teeth of the pair of sprockets 27a, 27b. 
On the other hand, the input shaft 1a can be connected to the ring sear 11 through another transmission shaft 13, which is disposed concentrically with the input shaft 1a and serves as a second transmission mechanism, in such a manner that the rotational power can be transmitted between them. Between this transmission shaft 13 and input shaft 1a, there is interposed the high speed clutch 14 in such a manner that it is arranged in series with the two shafts 13 and 1a. Therefore, the transmission 13, when the high speed clutch 14 is connected, can be rotated together with the input shaft 1a in the same direction and at the same speed.
Also, the present continuously variable transmission apparatus comprises the high speed clutch 14 and a low speed clutch 31 interposed between the outer peripheral edge portion of the carrier 24 and the axial-direction one end portion (in FIG. 12, the right end portion) of the ring gear 11. The low speed clutch 31 and high speed clutch 14 are structured such that, in case where any one of these clutches is connected, the connection of the other clutch is cut off. Also, in the example shown in FIG. 12, between the ring gear 11 and the fixed portion of the present continuously variable transmission apparatus such as a housing (not shown), there is interposed a reversing clutch 32. In case where any one of the low speed clutch 31 and high speed clutch 14 is connected, the connection of the reversing clutch 32 is cut off. Also, in case where the reversing clutch 32 is connected, the low speed clutch 31 and high speed clutch 14 are both disconnected.
In the case of the above structured continuously variable transmission apparatus, firstly, during the low speed running operation, the low speed clutch 31 is connected and, at the same time, the high speed clutch 14 and reversing clutch 32 are disconnected respectively. In this state, in case where the start clutch 17 is connected and the input shaft 1a is rotated, only the toroidal-type continuously variable transmission 9 is allowed to transmit the power from the input shaft 1a to the output shaft 20. In such low speed running operation, the transmission ratios between respective pairs of input side disks 3, 3 and output side disks 5, 5 are adjusted similarly to the case previously shown in FIGS. 10 and 11 in which only the toroidal-type continuously variable transmission 9 is allowed to transmit the power.
On the other hand, in the high speed running operation, he high speed clutch 14 is connected and, at the same time, the connection of the low speed clutch 31 and reversing clutch 32 is cut off. In this state, in case where the start clutch 17 is connected and the input shaft la is rotated, the transmission shaft 13 and planetary gear mechanism 10 are allowed to transmit the power from the input shaft 1a to the output shaft 20. That is, in case where the input shaft 1a is rotated in the high speed running operation, the rotational power of the input shaft 1a is transmitted through the high speed clutch 14 and transmission shaft 13 to the ring gear 11. And, the rotational power of the ring gear 11 is transmitted through the plurality of planetary gear sets 23, 23 to the sun gear 21, thereby rotating the output shaft 20 to which the sun gear 21 is fixed. In, this state, in case where the transmission ratio of the toroidal-type continuously variable transmission 9 is changed to thereby vary the revolving speeds (the rotation speeds around the periphery of the sun gear 21) of the respective planetary gear sets 23, 23, the transmission ratio of the whole of the continuously variable transmission apparatus can be adjusted.
In other words, in case where the input shaft 1a is rotated in the high speed running operation, the rotational power of the input shaft 1a is transmitted through the transmission shaft 13 and support plate 12 to the ring gear 11, thereby rotating the ring gear 11. And, the rotational power of the ring gear 11 is then transmitted through the plurality of planetary gear sets 23, 23 to the sun gear 21, thereby rotating the output shaft 20 to which the sun gear 21 is fixed. When the ring gear 11 serves as the input side, assuming that the respective planetary gear sets 23, 23 are not rotating (that is, they are not revolving around the periphery of the sun gear 21), the planetary gear mechanism 10 increases the revolving speed between the input shaft 1a and the output shaft 20 by the transmission ratio that corresponds to the ratio of the numbers of the ring gear 11 and sun gear 21. However, the respective planetary gear sets 23, 23 actually rotate around the periphery of the sun gear 21, while the transmission ratio of the whole of the continuously variable transmission apparatus varies according to the speed of the rotation (around the periphery of the sun gear 21) of the respective planetary gear sets 23, 23. Thus, in case where the transmission ratio of the toroidal-type continuously variable transmission 9 is changed to thereby change the revolving speed (the rotation speed around the periphery of the sun gear 21) of the respective-planetary gear sets 23, 23, the transmission ratio of the whole of the continuously variable transmission apparatus can be adjusted.
That is, in the above-mentioned high speed running operation, the respective planetary gear sets 23, 23 rotate around the periphery of the sun gear 21 in the same direction of the ring gear 11. And, the slower the rotation speed (around the periphery of the sun gear 21) of the respective planetary gear sets 23, 23 is, the faster the rotation speed of the output gear 20 with the sun gear 21 fixed thereto is. For example, in case where the rotation speed (around the periphery of the sun gear 21) of the respective planetary gear sets 23, 23 is equal to the rotation speed of the ring gear 11 (both of which are angular speeds), the rotation speed of the ring gear, 11 is equal to that of the output shaft 20. On the other hand, in case where the rotation speed (around the periphery of the sun gear 21) of the respective planetary gear sets 23, 23 is faster than the rotation speed of the ring gear 11, the rotation speed of the output shaft 20 is slower than that of the ring gear 11.
Therefore, in the high speed running operation, as the transmission ratio of the toroidal-type continuously variable transmission 9 is changed toward the speed reducing side, the transmission ratio of the whole of the continuously variable transmission apparatus is changed toward the speed increasing side. In the state of such high speed running operation, to the toroidal-type continuously variable transmission 9, there is applied torque not from the input side disk 3 but from the output side disk 5 (that is, assuming that torque to be applied in the low speed running operation is positive, negative torque is applied). In other words, in a state where the high speed clutch 14 is connected, the torque, which has been transmitted from the engine 15 to the input shaft 1a, is then transmitted through the transmission shaft 13 to the ring gear 11 of the planetary gear mechanism 10. Therefore, there remains little torque that is to be transmitted from the input shaft 1a side to the input side disks 3, 3 constituting the toroidal-type continuously variable transmission 9.
On the other hand, part of the torque transmitted through the transmission shaft 13 to the ring gear 11 is transmitted from the respective planetary gear sets 23, 23 to the respective output side disks 5, 5 through the carrier 24 and first power transmission mechanism 25. The torque to be applied from the output side disks 5, 5 to the toroidal-type continuously variable transmission 9 reduces as the transmission ratio of the toroidal-type continuously variable transmission 9 is varied toward the speed reducing side in order to change the whole of the continuously variable transmission apparatus toward the speed increasing side. As a result of this, in the high speed running operation, by reducing the torque to be input to the toroidal-type continuously variable transmission 9, the durability of the composing parts of the toroidal-type continuously variable transmission 9 can be enhanced.
Further, when rotating the output shaft 20 reversely so as to back the vehicle, the low speed and high speed clutches 31, 14 are disconnected and, at the same time, the reversing clutch 32 is connected. As a result of this, the ring gear 11 is fixed; and, the planetary gear sets 23, 23, while they are being in meshing engagement with the ring gear 11 and sun gear 21, are rotated around the periphery of the sun gear 21. And, the sun gear 21 and output shaft 20 with the sun gear 21 fixed thereto are rotated in the opposite direction to the above-mentioned low speed and high speed running operations.
Next, FIGS. 13 and 14 show a more concrete example of the continuously variable transmission apparatus shown in the above-mentioned FIG. 12. By the way, the inventors have conducted a series of tests to be discussed below using the continuously variable transmission apparatus shown in FIGS. 13 and 14, which have-conducted us to development of the present invention. The present continuously variable transmission apparatus comprises an input shaft 1b, an output shaft 20a, a toroidal-type continuously variable transmission 9a, a planetary gear mechanism 10a, a first power transmission mechanism 25a, and a transmission shaft 13a constituting a second-power transmission mechanism. Of the above composing parts, the input shaft 1b is connected to a drive source such as an engine 15 (see FIG. 12) and can be driven or rotated by the drive source. Also, the output shaft 20a, which is used to take out the power based on the rotational movement of the input shaft 1b, is connected through a differential gear (not shown) to a wheel drive shaft.
Also, the toroidal-type continuously variable transmission 9a is of a double cavity type and includes trunnions 7, 7 and power rollers 4, 4 each by threes in each of the two cavities, that is, a total of six trunnions 7 and power rollers 4. In order to construct such toroidal-type continuously variable transmission 9a, a pair of input side disks 3, 3 are respectively supported on the two end portions of the input shaft 1b in such a manner that they can be rotated in synchronization with the input shaft 1b and the respective inner surfaces 3a, 3a of the pair of input side disks 3, 3 are disposed -opposed to each other. Of the two input side disks 3, 3, the input side disk 3 situated on the base end side (that is, on the drive source side; and, in FIG. 13, on the left side) of the input shaft 1b is supported on the input shaft 1b through a ball spline 33 in such a manner that it can be shifted in the axial direction of the input shaft 1b. On the other hand, the input side disk 3 on the leading end side (that is, on the side far from the drive source; and, in FIG. 13, on the right side) of the input shaft 1b is fixed to the input shaft 1b in such a manner that the back surface of the input side disk 3 is held by a loading nut 34 while the input side disk 3 is spline engaged with the leading end portion of the input shaft 1b. 
And, on the portions of the peripheries of the middle portions of the input shaft 1b that exist between the pair of input side disks 3, 3, there are disposed a pair of output side disks 5, 5 in such a manner that the inner surfaces 5a, 5a of the output side disks 5, 5 are respectively opposed to the inner surfaces 3a, 3a of the input side disks 3, 3 and the output side disks 5, 5 can be respectively rotated in synchronization with their associated input side disks 3, 3. And, the power rollers 4, 4, which are rotatably supported on the inner surfaces of their associated trunnions 7, 7, are held by and between the inner surfaces 3a, 3a of the input side disks 3, 3 and the inner surfaces 5a, 5a of the output side disks 5, 5.
In order to support these trunnions 7, 7, a frame 37 is connected and fixed to a mounting portion 36 formed in the inner surface of a casing 35 by three studs 39, 39 respectively inserted through their associated mounting holes 38, 38 formed at three positions in the outside diameter side end portion of the frame 37 and three nuts 40, 40 respectively threadedly engaged with their associated studs 39, 39. In the illustrated example, a gear housing 41 is fixed between the mounting portion 36 and frame 37 by these studs 39, 39 and nuts 40, 40. On the inside diameter side of the gear housing 41, an output sleeve 42 with the two end portions thereof unevenly (projectingly and recessedly) engaged with the pair of output side disks 5, 5 is rotatably supported by a pair of rolling bearings 43, 43; and, an output gear 44 disposed on the outer peripheral surface of the middle portion of the output sleeve 42 is stored in the interior of the gear housing 41.
Also, the frame 37 is formed in a star shape as a whole, while the diameter-direction middle portion or outside diameter side portion of the frame 37 is forked to thereby provide three hold portions 45, 45 at regular intervals in the circumferential direction of the frame 37. And, the middle portions of three support pieces 46, 46 are respectively pivotally supported on the diameter-direction middle portions of their associated hold portions 45, 45 by their associated second pivot shafts 47, 47. Each of the three support pieces 46, 46 is composed of a cylinder-shaped mounting portion 48 to be disposed on the periphery of its associated one of the second pivot shafts 47, 47, and a pair of support plate portions 49, 49 which are respectively provided on and projected from the outer peripheral surface of the mounting portion 48 outwardly in the diameter direction. An angle of intersection between the pair of support plate portions 49, 49 is set an angle of 120°. Therefore, the support plate portions 49, 49, which adjoin each other in the circumferential direction of the frame 37, are parallel to each other.
In these support plate portions 49, 49, there are formed circular holes 50, 50, respectively. When the support-pieces 46, 46, are respectively held at their neutral positions, the circular holes 50, 50, which are formed in the support plate portions 49, 49 of the support pieces 46, 46 adjoining each other in the circumferential direction of the frame 37, are concentric with each other. And, pivot shafts 8, 8, which are respectively disposed on the two end portions of each of the trunnions 7, 7, are respectively supported within their associated circular holes 50, 50 by their associated radial needle roller bearings 51, 51. The outer peripheral surfaces of outer rings 52, 52, which respectively constitute their associated radial needle roller bearings 51, 51, are respectively formed as a spherical-shaped convex surface. The outer rings 52, 52 are respectively inserted into their associated circular holes 50, 50 in such a manner that they can be prevented against shaky motion and can be swingly shifted. Also, in the support plate portions 49, 49, there are formed screw holes 53, 53 respectively; and the spherically-convex-surface-shaped leading end faces of studs 54, 54 threadedly engaged into the screw holes 53, 53 are respectively butted against the two end faces of the respective trunnions 7, 7. Thanks to this structure, the shift amounts of the respective trunnions 7, 7 with respect to the circumferential direction with the input shaft 1b as the center thereof can be mechanically synchronized with each other.
On the inner surfaces of the trunnions 7, 7 supported in the interior of the casing 35 in the above-mentioned manner, there are supported their associated power rollers 4, 4 through shift shafts 55 structured such that the base half portions and front half portions thereof are formed eccentric to each other. Also, between the outer end faces of the power rollers 4, 4 and the inner surfaces of the trunnions 7, 7, there are respectively interposed thrust ball bearings 56 and thrust needle roller bearings 57 in the order starting from the power roller 4 side. Of these bearings, the thrust ball bearings 56 not only support thrust loads to be applied to the power rollers 4 but also allow the power rollers 4 to rotate. On the other hand, in the case of the thrust needle roller bearings 57, in case where the composing parts of the toroidal-type continuously variable transmission 9a are elastically deformed when the toroidal-type continuously variable transmission 9a is in operation and thus the shift shafts 55 are swung about their respective base half portions to cause the power rollers 4 to shift in the axial direction of the input shaft 1b, the thrust needle roller bearings 57 allow such shifting motion of the power rollers 4 to be executed smoothly.
The peripheral surfaces 4a, 4a of the power rollers 4, 4 supported on the inner surfaces of the trunnions 7, 7 in the above-mentioned manner are contacted with the inner surfaces 3a, 3a of the input side and output side disks 3, 5. Also, an oil pressure type of pressing device 18a is assembled between the input shaft 1b and the input side disk 3 on the base end side of the input shaft 1b to thereby secure the surface pressure of the contact portions (traction portions) between the respective surfaces 4a, 3a, 5a, so that the power transmission by the toroidal-type continuously variable transmission 9a can be carried out with high efficiency.
In order to construct the pressing device 18a, an outwardly facing flange portion 58 is fixedly disposed on the near-to-base-end portion of the outer peripheral surface of the input shaft 1b and, at the same time, a cylinder 59 is oiltight fitted with and supported by the outer surface of the above-mentioned base-end-side input side disk 3 in such a manner it projects axially from the outer surface (in FIG. 13, the left surface) of the present input side disk 3. The inside diameter of the cylinder 59 is small in the axial-direction middle portion thereof and large in the two end portions thereof; and, the present input side disk 3 is oiltight fitted with the inner surface of the large-diameter portion of the leading end side of the cylinder 59 in such a manner that it can be shifted in the axial direction. Also, in the inner peripheral surface of the middle portion of the cylinder 59, there is formed an inwardly-facing-flange-shaped partition plate portion 60. Further, between the inner peripheral surface of the cylinder 59 and the outer peripheral surface of the input shaft 1b, there is interposed a first piston member 61.
The first piston member 61 includes an outwardly-facing-flange-shaped partition plate 63 formed on the outer peripheral surface of the middle portion of a support tube portion 62 which can be fitted with the outer surface of the input shaft 1b; and, the outer peripheral edge of the partition plate 63 is oiltight slidingly contacted with the small-diameter portion of the inner peripheral surface of the cylinder 59 in such a manner that it can be shifted in the axial direction. Also, in this state, the inner peripheral edge of partition plate portion 60 is oiltight slidingly contacted with the outer peripheral surface of the support tube portion 62 in such a manner that it can be shifted in the axial direction. Further, between the outer peripheral surface of the base end portion of the support tube portion 62 and the inner peripheral surface of the base end portion of the cylinder 59, there is interposed a circular-ring-shaped second piston member 64. The second piston member 64, when the base-end-side side surface thereof is contacted with the flange portion 58, can be prevented from shifting in the axial direction and, at the same time, can keep oiltight between the inner and outer peripheral edges thereof, the base end portion outer peripheral surface of the support tube portion 62 and the base end portion inner peripheral surface of the cylinder 59.
Also, the cylinder 59 including the partition plate portion 60 is pressed toward the input side disk 3 by a preload spring such as a countersunk plate spring 65 which is interposed between the partition plate portion 60 and second piston member 64. Therefore, the present-input side disk 3 is pressed at least (that is, even in a state where pressure oil is not introduced in the interior of the pressing device 18a) by a pressing force corresponding to the elasticity of the countersunk spring 65, so that the input side disk 3 applies the surface pressure corresponding to such elasticity to the contact portions between the respective surfaces 4a, 3a, 5a. In this case, this elasticity is restricted to such a degree that, when a very small level of power is transmitted by the toroidal-type continuously variable transmission 9a, slippage (excluding spin which is unavoidable) can be prevented from occurring in the respective contact portions between the respective surfaces 4a, 3a, 5a. 
Also, the oil pressure can be introduced through a center hole 66 formed in the input shaft 1b into oil pressure chambers respectively existing between the second piston member 64 and partition plate portion 60 as well as between the partition wall plate 63 and input side disk 3. This center hole 66 communicates through an oil control valve (not shown) with an oil pressure source such as a pressurizing pump (not shown) When the continuously variable transmission apparatus including the toroidal-type continuously variable transmission 9a is in operation, the oil pressure, which is adjusted by the oil pressure control valve in accordance with the size of the power to be transmitted, is introduced into the respective oil pressure chambers to thereby press the input side disk 3, so that the surface pressure corresponding to the size of the above power is applied to the respective contact portions between the respective surfaces 4a, 3a, 5a. 
Also, transmission of the rotational power to the input shaft 1b from a drive shaft 67 communicating with a drive source such as an engine is carried out through the flange portion 58. For this purpose, at a plurality of portions in the outer peripheral edge portion of the flange portion 58, there are formed notches 68, 68; and, these notches 68, 68 are respectively engaged with driving projecting portions 69, 69 formed in the end portion of the drive shaft 67. Also, for the above purpose, in the case of the present apparatus, an outwardly-facing-flange-shaped connecting portion 70 is formed in the end portion of the drive shaft 67, while the driving projecting portions 69, 69 are formed in the near-to-outside-diameter end portion of one surface of the connecting portion 70.
Further, actuators 71a, 71b each of an oil pressure type are attached to the respective trunnions 7, 7 so that the trunnions 7, 7 can be driven or shifted in the axial directions of their associated pivot shaft 8, 8 respectively disposed on the two end portions of each of the trunnions 7. Of the trunnions 7, the trunnion 7 disposed in the central portion of the lower side of FIG. 14 can be driven or shifted through lever arms 72, 72 in the axial directions of the pivot shafts 8, 8 disposed on the two end portions thereof by a pair of actuators 71a, 71a which are each of a single action type (that is, a type which is capable of obtaining only the push-out direction force) and the pressing directions of which are opposite to each other. When shifting each of the present trunnion 7, the pressure oil is fed only into the oil pressure chamber of one of the actuators 71a, whereas the oil pressure chamber of the other actuator 71a is set free. On the other hand, trunnions 7, 7 disposed on the two sides of the upper portion of FIG. 14 can be driven or shifted in the axial directions of the pivot shafts 8, 8 disposed on the two end portions of the trunnions 7, 7 by actuators 71b, 71b each of a double action type (a type which is capable of obtaining the push-out direction force or the pull-in direction force in accordance with switching of the supply direction of the pressure oil).
A total of six trunnions 7, 7, which are disposed in the toroidal-type continuously variable transmission 9a, may be shifted by the same length in synchronization with each other by supplying the same amount of pressure oil to the actuators 71a, 71b using a control valve. For this purpose, to the end portion of a rod 73 which can be shifted together with any one (in the illustrated example, in FIG. 14, the trunnion 7 on the upper left side) of the trunnions 7, there is fixed a precess cam 74 and the attitude of this trunnion 7 can be transmitted through a link arm 75 to a spool 76 of the above control valve.
The planetary gear mechanism 10a, which is combined with the above-structured toroidal-type continuously variable transmission 9a, comprises a sun gear 21a, a ring gear 11a, and planetary gear sets 23a, 23a. The sun gear 21a is fixed to the input side end portion (in FIG. 13, the left end portion) of the output shaft 20a. Therefore, the output shaft 20a can be rotated as the sun gear 21a is rotated. On the periphery of the sun gear 21a, there is supported the ring gear 11a in such a manner that it is concentric with the sun gear 21a and can be rotated. And, between the inner peripheral surface of the ring gear 11a and the outer peripheral surface of the sun gear 21a, there are interposed a plurality of planetary gear sets 23a, 23a each set consisting of a pair of planetary gears 22a, 22b combined together. Each pair of planetary gears 22a, 22b are meshingly engaged with each other, the planetary gear 22a disposed on the outside diameter side is meshingly engaged with the ring gear 11a, and the planetary gear 22b disposed on the inside diameter side is meshingly engaged with the sun gear 21a. The thus structured planetary gear sets 23a, 23a are rotatably supported on one side surface (in FIG. 13, the left side surface) of a carrier 24a. And, the carrier 24a is rotatably supported on the periphery of the middle portion of the output shaft 20a. 
Also, the carrier 24a is connected to a pair of output disks 5, 5 structuring the toroidal-type continuously variable transmission 9a by the first power transmission mechanism 25a in such a manner that the rotational power can be transmitted between them. In order to constitute the first power transmission mechanism 25 at there is disposed a transmission shaft 26a which extends in parallel to the input shaft 1b and the output shaft 20a, and a gear 77 fixed to the one-end portion (in FIG. 13, the left end portion) of the transmission shaft 26a. Also, on the periphery of the middle portion of the output shaft 20a, there is disposed a sleeve 78 in such a manner that it can be rotated; and, a gear 79 supported on the outer peripheral surface of the sleeve 78 is meshingly engaged with a gear 80 fixedly secured to the other end portion (in FIG. 13, the right end portion) of the transmission shaft 26a through an idler gear (not shown). Further, on the periphery of the sleeve 78, there is supported the carrier 24a through a circular-ring-shaped connecting bracket 81 in such a manner that it can be rotated in synchronization with the sleeve 78. Therefore, as the output side disks 5, 5 are rotated, the carrier 24a can be rotated in the opposite direction to the output side disks 5, 5 at the speed that corresponds to the numbers of teeth of the respective gears 44, 77, 79, 80. Also, between the carrier 24a and output shaft 20a, there is interposed a low speed clutch 31a. 
On the other hand, the input shaft 1b and ring gear 11a are connected to each other through the input side disk 3 supported on the leading end portion of the input shaft 1b and the transmission shaft 13a disposed concentrically with the input shaft 1b in such a manner that the rotational power can be transmitted between them. To attain this, a plurality of projecting portions 82, 82 are provided on and projected from the portion of the outer surface (in FIG. 13, the right side surface) of the present input side disk 3, that is, the half portion of the present outer surface that is situated nearer to the outside diameter of the input side disk 3 than the central portion of the present outer surface with respect to the diameter direction of the input side disk 3. In the case of the present example, these projecting portions 82, 82 are respectively formed in an arc shape and are arranged intermittently and at regular intervals on the same arc with the axis of the input side disk 3 as the center thereof. And, the portions, which exist between the circumferential-direction end faces of the projecting portions 82, 82 adjoining each other in the circumferential direction of the input side disk 3, are formed as securing notches 83, 83.
On the other hand, on the base end portion of the transmission shaft 13a, there is disposed a transmission flange 85 through a conically-cylindrical-shaped transmission cylinder portion 84. And, in the buter peripheral edge portion of the transmission flange 85, there are disposed the same number of transmission projecting pieces 86, 86 as the number of securing notches 83, 83 at regular intervals with respect to the circumferential direction of the transmission flange 85. And, the transmission projecting pieces 86, 86 are respectively engaged with their associated securing notches 83, 83, thereby allowing torque to be transmitted between the input side disk 3 and transmission shaft 13a. Since the diameter of the mutually engaged portions between the transmission projecting pieces 86, 86 and securing notches 83, 83 is sufficiently large, sufficiently large torque can be transmitted between the input side disk 3 and transmission shaft 13a. 
By the way, in order that the torque to be transmitted between the input side disk 3 and transmission shaft 13a can be increased as much as possible, preferably, the projecting portions 82, 82 may be provided on the near-to-outside-diameter end portion (outer peripheral edge portion) of the outer surface of the input side disk 3. However, in the case of the present example, in order to secure the finishing precision of the inner surface 3a of the input side disk 3, a flat portion 87 is formed in such portion of the outer surface of the input side disk 3 that is situated nearer to the outside diameter of the input side disk 3 than the projecting portions 82, 82, so that, when finishing the inner surface 3a of the input side disk 3, the near-to-outside-diameter portion of the outer surface of the input side disk 3 can be supported using the flat portion 87. Also, the transmission projecting pieces 86, 86 are formed such that they can be engaged into their associated securing notches 83, 83 with no shaky motion between them.
Also, the leading end portion (in FIG. 13, the right end portion) of the transmission shaft 13a is rotatably supported on the center portion of the sun gear 21a. Further, the ring gear 11a is supported on the periphery of the middle portion of the transmission shaft 13a through a circular-ring-shaped connecting bracket 88 spline engaged with the transmission shaft 13a and through a high speed clutch 14a (which will be discussed later) in such a manner that the ring gear 11a can be rotated in synchronization with the transmission shaft 13a. Therefore, while the high speed clutch 14 is connected, as the input shaft 1b is rotated, the ring gear 11a can be rotated together with the input shaft 1b in the same direction and at the same speed as the input shaft 1b. 
Further, between the ring gear 11a and the fixed-portion of the casing 35 such as a fixed wall 89 formed within the casing 35, there is interposed a reversing clutch 32a. The reversing clutch 32a, high speed clutch 14a and low speed clutch 31a are all wet-type multiple disk clutches each comprising a plurality of friction plates and a plurality of separate plates which are arranged alternately. These clutches 32a, 14a and 31a can be respectively connected or disconnected in accordance as the pressure oil is supplied into a high speed oil pressure cylinder 90, a low speed oil pressure cylinder 91, and a reversing oil pressure cylinder 92 which are respectively attached to the clutches 32a, 14a and 31a. Also, in case where any one of these clutches is connected, the connection of the remaining two clutches is cut off. By the way, while the clutches 32a, 14a and 31a are equal in the effective radius to each other, the friction plates and separate plates are different in number from each other. That is, the number of the plates constituting the low speed clutch 31a required to transmit the largest torque is set the largest (for example, eight plates each); and, the number of the plates constituting the reversing clutch 32a and high speed clutch 31a required to transmit relatively small torque is set smaller (for example, five plates each) than the number of the plates constituting the low speed clutch 31a. 
The above-structured continuously variable transmission apparatus is similar in operation to the conventional structure shown in the previously discussed FIG. 12. That is, firstly, in the low speed running operation (low speed mode), the oil pressure is introduced into the low speed oil pressure cylinder 91 to thereby connect the low speed clutch 31a and, at the same time, oil pressures existing within the high speed and reversing oil pressure cylinders 90, 92 are discharged therefrom to thereby cut off the connection of the high speed clutch 14a and reversing clutch 32a. 
Also, in the high speed running operation (high speed mode), the oil pressure is introduced into the high speed oil pressure cylinder 90 to thereby connect the high speed clutch 14a and, at the same time, oil pressures within the low speed and reversing oil pressure cylinders 91, 92 are discharged therefrom to thereby cut off the connection of the low speed clutch 31a and reversing clutch 32a. In this state, in case where the input shaft 1b is rotated, the transmission shaft 13a serving as the second power transmission mechanism and the planetary gear mechanism 10a transmit the rotational power from the input shaft 1b to the output shaft 20a. In this state, by changing the transmission ratio of the toroidal-type continuously variable transmission 9a to thereby change the revolving speeds (around the sun gear 21a) of the respective planetary gear sets 23a, 23a, the transmission ratio of the whole of the continuously variable transmission apparatus can be adjusted.
Further, when reversing the output shaft 20a in order to back the vehicle, the oil pressures within the low speed and high speed-oil pressure cylinders 90, 91 are discharged therefrom to thereby cut off the connection of the low speed and high speed clutches 31a, 14a and, at the same time, the oil pressure is introduced into the reversing oil pressure cylinder 92 to thereby connect the reversing clutch 32a. In this state, the sun gear 21a and the output shaft 20a fixed to the sun gear 21a are rotated in the opposite direction to the direction in the previously described low speed and high speed running operations.
In the case of the conventional continuously variable transmission apparatus structured and operated in the above-mentioned manner, according to a study made by the present inventors, it has been found that, in switching the low speed and high speed modes over to each other, the number of revolutions of the engine can vary suddenly to thereby raise a possibility of giving a driver an incongruous feeling. Also, according to the study made by the present inventors, it has also been found that such sudden variation in the number of revolutions of the engine is caused by the following facts: that is, in the above mode switching time, there exists a moment when the connection of the low speed and high speed clutches 31a, 14a (of course, the connection of the reversing clutch 32a as well) is cut off; and, the low speed and high speed clutches 31a, 14a are different in capacity from each other.
Now, description will be given below of this point with reference to FIGS. 15 to 17 in addition to FIG. 13.
Of the above figures, FIG. 15 shows an apparatus used in a test which was conducted to know not only the timing for signaling for instructing the connection or disconnection of the low speed and high speed clutches 31a, 14a (in FIGS. 15, 31, 14) but also the connecting states of these two clutches 31a, 14a based on this signal. The present apparatus, using the continuously variable transmission apparatus previously shown in FIGS. 13 and 14, is used to find not only the timing for issuance of the signal for instructing the connection or disconnection of the high speed and low speed clutches 14a, 31a, but also the timing at which these two clutches 14a, 31a are actually connected and disconnected.
By the way, the timing for connection and disconnection of the two clutches 14a, 31a was judged according to the oil pressures of the high speed and low speed oil pressure cylinders 90, 91 which are respectively attached to the two clutches 14a, 31a. That is, in case where the oil pressures of the high speed and low speed oil pressure cylinders 90, 91 attached to the two clutches 14a, 31a are low, it was judged that pressing pistons respectively disposed within these cylinders are movable, there exist gaps respectively between the friction plates and separate plates constituting these clutches, and the connection of the present clutches is cut off. On the other hand, in case where the oil pressures of the high speed and low speed oil pressure cylinders 90, 91 attached to the two clutches 14a, 31a are sufficiently high, it was judged that pressing pistons respectively disposed within these cylinders cannot be moved, the friction plates and separate plates constituting these clutches are contacted with each other, and the present clutches are connected. Further, in case where the oil pressures of the high speed and low speed oil pressure cylinders 90, 91 attached to the two clutches 14a, 31a are intermediate values, it was judged that the present clutches are held in a so called clutch slipping state in which they transmit the rotational power while slipping.
Firstly, description will be given below of the test apparatus shown in FIG. 15 that is structured based on the structure of an actually used continuously variable transmission apparatus. By the way, a toroidal-type continuously variable transmission 9 and a planetary gear mechanism 10 shown in FIG. 15 are similar to those previously shown in FIG. 12; and, a first power transmission mechanism 25 is also structured such that sprockets 27a, 27b and a chain 28 are incorporated therein. With regard to reference characters, there are used the same reference characters as in FIG. 12 and thus the description thereof is omitted here. By the way, although there is shown a start clutch 17 in FIG. 15, in the case of an actual test apparatus, this start clutch 17 is omitted; and, a drive source and the input shaft 1a of the toroidal-type continuously variable transmission 9 are connected directly to each other. While a reversing clutch 32 is disposed, an oil pressure pipe is omitted; and, therefore, this reversing clutch 32 does not function.
As pressure oil for connecting together the high speed and low speed clutches 14, 31, there is used pressure oil which is sucked out from an oil tank 93 (in actual assembly to a vehicle, an oil pan) and is then jetted out from a pressurizing pump 94. Between the pressurizing pump 94 and the above-mentioned high speed clutch 14 {specifically, the high speed oil pressure cylinder 90 (FIG. 13) for connecting and disconnecting the high speed clutch 14}, there is interposed a high speed side switch valve 95; and, similarly, between the pressurizing pump 94 and the above-mentioned low speed clutch 31 {specifically, the low speed oil pressure cylinder 91 (FIG. 13) for connecting and disconnecting the low speed clutch 31}, there is interposed a low speed side switch valve 96. These switch valves 95 and 96 respectively turn on/off solenoids attached thereto in accordance with a signal from a controller 97 to thereby switch the following two modes over to each other: that is, one mode in which the oil tanks 14, 31 are allowed to communicate with the jet-out port of the pressurizing pump 94; and, the other mode in which the oil tanks 14, 31 are allowed to communicate with the oil tank 93. Also, on such portions of the oil pressure pipe that exist between the switch valves 95, 96 and clutches 14, 31, there are mounted high speed side and low speed side pressure gauges 98, 99 respectively, so that the oil pressures of the clutches 14, 31 (specifically, the high speed and low speed oil pressure cylinders 90, 91 for connecting and disconnecting the clutches 14, 31) can be measured.
The present inventors, using the above-structured test apparatus, measured not only the timing at which a signal for switching the high speed side and low speed side switch valves 95, 96 is issued by the controller 97 but also the timing at which the high speed and low speed clutches 14, 31 are actually connected. The results of the measurement are shown in FIGS. 16A to 17B. FIGS. 16A to 17B respectively show not only variations in an instruction signal given to the high speed and low speed clutches 14, 31 in the switching time of the low speed and high speed modes but also variations in the connecting states of the two clutches 31, 14. Also, of the two FIGS. 16A to 17B, FIGS. 16A and 16B show the variations which occur when switching the low speed mode over to the high speed mode, while FIGS. 17A and 17B show the variations occurring when switching the high speed mode over to the low speed mode.
Also, the horizontal axes of FIGS. 16A to 17B express the elapsed time; and, the vertical axes of FIGS. 16A and 17A express the above-mentioned instruction signals, whereas the vertical axes of FIGS. 16B and 17B express the connecting states of the clutches. By the way, these instruction signals are instruction signals which are transmitted from the controller 97 to solenoids attached to the respective switch valves 96, 95 in order to control the supply of the pressure oil to the low speed and high speed oil pressure cylinders 91, 90 for connecting and disconnecting the clutches 31, 14; and, the vertical axes of FIGS. 16A and 17A show the voltages of these instruction signals. In case where these voltage are positive, the clutches are connected, and, in case where they are negative, the connection of the clutches is cut off. Also, the connecting states of the clutches are expressed by the measured values of the pressure gauges 99, 98 that are proportional to the contact pressure ratio between the friction plates and separate plates constituting the two clutches 31, 14.
Further, solid lines shown in FIGS. 16A to 17B express the above-mentioned variations with respect to the low speed clutch 31, whereas broken lines express the variations with respect to the high speed clutch 14, respectively.
In the test, when switching the low speed mode and high speed mode over to each other, in both of switching directions, the instruction signals to be applied to the switch valves 96, 95, as shown in FIGS. 16A and 17A, were switched within the time of about 0.2 sec. That is, in about 0.2 sec. after issuance of a signal for cutting off the connection of the currently connected clutch, there was issued a signal for connecting the currently disconnected clutch. As the results of the test conducted under these conditions, it has been found that, when switching the low speed mode and high speed mode over to each other, in case where the timings for issuance of the instruction signals are set the same regardless of the switching directions, a continuous time, during which neither of the clutches are not connected, increases.
That is, as can be seen obviously from FIGS. 16B and 17B, in the case of the clutch to which a signal for cutting off the connection thereof has been transmitted, the connection of the clutch is cut off in a very short time (see the solid line shown in FIG. 16B and the broken line shown in FIG. 17B). On the other hand, as can also be seen obviously from FIGS. 16B and 17B, in the case of the clutch to which a signal for connection has been transmitted, there is generated a slight time delay between reception of the signal and completion of the connection of the of the clutch. Also, as can be understood clearly from comparison between the broken line shown in FIG. 16B and the solid line shown in FIG. 17B, the degree of the time delay varies according to the mode switching directions.
Specifically, the time necessary for connection of the currently disconnected clutch (the time during which the clutch is held in the clutch-slipping state) is longer in the switching operation from the high speed mode to the low speed mode shown in FIG. 17B than in the switching operation from the low speed mode to the high speed mode shown in FIG. 16B.
The present inventors not only have studied the reason for generation of the slight time delay between reception of the connection signal by the clutch and completion of the actual connection of the clutch and the reason for the difference of the time delay according to the mode switching directions, but also have repeatedly conducted the test for confirmation of these reasons. Our study and confirmation test results have found the following facts.
Firstly, the time delay is caused by the fact that it takes some time to complete the full-stroke movements of the pistons incorporated in the high speed and low speed oil pressure cylinders 90, 91. That is, to connect together the friction plates and separate plates constituting the clutches 14, 31 in order to connect the clutches 14, 31, it is necessary to shift the respective friction and separate plates as well as the pressurizing pistons that are incorporated in the oil pressure cylinders 90, 91. This shifting operation is carried out by introducing pressure oil into the oil pressure cylinders 90, 91; however, due to resistance within the pipes used to introduce the pressure oil, it inevitably takes time to complete introduction of a sufficient quantity of pressure oil. This gives rise to the above-mentioned generation of the time delay.
Next the reason for the different degrees of the time delay according to the mode switching directions is that, between the high speed clutch 14 and low speed clutch 31, the strokes of the pressurizing pistons incorporated in the oil pressure cylinders 90, 91, which are necessary to shift the pressurizing pistons up to such positions as to allow completion of connection of the high speed clutch 14 and low speed clutch 31, are different.
That is, as described before, in the case of the low speed clutch 31 which is required to transmit large torque, the number of the plates is large; and, on the other hand, in the case of the high speed clutch 14 which is required to transmit relatively small torque, the number of the plates is small. As the number of the plates increases, the strokes of the pressurizing pistons incorporated in the oil pressure cylinders 90, 91, which are necessary to bring the disconnected clutches into connected states, increase. For this reason, as shown by the broken line in FIG. 16B and by the solid line in FIG. 17B, the time delay in the switching operation from the high speed mode to the low speed mode (the solid line in FIG. 17B) is larger than the time delay in the switching operation from the low speed mode to the high speed mode (the broken line in FIG. 16B).
During the above time delay, since the engine serving as the power source is not connected (or is imperfectly connected) to the drive wheel, the power cannot be transmitted from the engine to the drive wheel. To confirm the behavior that occurs in this case, the present inventors have conducted a test for switching the low speed mode and high speed mode over to each other using a test apparatus shown in FIG. 15. In this test, the rotation speed of an input shaft 1a is fixed to 200min−1; and, any other rotary shaft was not connected to the end portion of an output shaft 20, while the output shaft 20 was set in such a manner that it can be rotated freely. Also, in a state where the input shaft 1a and output shaft 20 were the same in the speed (that is, the transmission ratio was 1), in order to switch the above two modes over to each other, the low speed clutch 31 and high speed clutch 14 were connected and disconnected. And, the present inventors measured variations in the rotation speed of the output shaft 20 occurring due to the switching of these two modes over to each other.
As a result of this test, not only in the switching operation from the low speed mode to the high speed mode but also in the switching operation from the high speed mode to the low speed mode, the rotation speed of the output shaft 20 was lowered down (to a value lower than 200min−1). Specifically, however, in the switching operation from the low speed mode to the high speed mode, the rotation speed of the output shaft 20 was lowered only by 20min−1 from 200min−1 to 180min−1; and, on other hand, in the switching operation from the high speed mode to the low speed mode, the rotation speed of the output shaft 20 was lowered no less than 80min−1 from 200min−1 to 120min−1. And, in either case, as the clutch was connected after the above-mentioned time delay, the rotation speed of the output shaft 20 returned back to 200min−1.
Since the above test was done in a state where the output shaft 20 can be rotated freely, in the mode switching operations, the rotation speed of the output shaft 20 was lowered. However, in an actual case, the output shaft 20 and the drive wheel are mechanically connected to each other through a propeller shaft and a differential gear. Therefore, in the mode switching operations, there is no possibility that the rotation speed of the output shaft 20 can be lowered but, in the clutch-slipping state, the rotation speed of the input shaft 1a increases. For this reason, in the mode switching operations, after the rotation speed of the engine serving as the drive source increases for an instant regardless of the operation of an accelerator pedal, the clutch is connected.
An increase in the rotation speed of the engine regardless of the operation of the accelerator pedal, even for an instant, is not preferable because it gives a driver an incongruous feeling. Also, in case where the clutch is connected after the rotation speed of the engine increases for an instant, there is a possibility that the whole of the power transmission system including the continuously variable transmission apparatus can be vibrated, which also gives the driver an incongruous feeling. There is a possibility that such incongruous feeling can increase especially in the switching operation from the high speed mode to the low speed mode.
The present invention aims at eliminating the above-mentioned drawbacks found in the conventional continuously variable transmission apparatus. Accordingly, it is an object of the present invention to provide a continuously variable transmission apparatus which can reduce the above-mentioned time delay at least in the switching operation from the high speed mode to the low speed mode.
Also, as another examples of a conventional toroidal-type continuously variable transmission, there are known toroidal-type continuously variable transmissions which are disclosed, for example, in JP-2734583 and JP-A-5-39850. In each of the toroidal-type continuously variable transmissions disclosed in these publications, between an input side disk and an output side disk, there are interposed a plurality of power rollers. The power rollers are respectively supported by their associated trunnions in such a manner that they can be swingly rotated. When changing the rotation speed between input and output shaft, the power rollers are swingly rotated to thereby change the rotation radius ratio of the contact point between the input side and output side disks.
In order to swing and rotate the power rollers, the trunnions are respectively moved by a desired amount in the axial directions of the respective trunnion shafts by their respective actuators including their respective oil pressure pistons to thereby offset the centers of rotation of the power rollers with respect to the centers of rotation of the input side and output side disks. In correspondence to the offset amounts of the trunnions, in the contact points between the input side and output side disks, there are generated moment forces which can swing and rotate the power rollers; and, due to such moment forces, the power rollers are swingly rotated at an angle which corresponds to the desired transmission ratio.
The above actuators move the trunnions in the axial direction thereof by the desired amount by driving their pistons using the pressure of oil which can be controlled by a transmission control valve. In order to stabilize the swingly rotational operations of the power rollers caused by the movements of the trunnions, for example, as disclosed in JP-A-11-294549, there is known a technique which feedbacks the transmission control valve the shift amounts of the trunnions (the sum totals of the swung-rotation-direction shift amounts of the trunnions and the swung rotation angular amounts of the trunnions) by a feedback mechanism using a precess cam.
When incorporating a toroidal-type continuously variable transmission into an actual vehicle, as disclosed in JP-A-10-196759, there is proposed a technique in which the toroidal-type continuously variable transmission is combined with a planetary gear mechanism. Here, FIG. 18 shows a continuously variable transmission apparatus which is referred to as a power split type of continuously variable transmission apparatus by the present inventors.
The present continuously variable transmission apparatus comprises a toroidal-type continuously variable transmission 102, a planetary gear mechanism 103, a first power transmission mechanism 104, and a second power transmission mechanism 105. The two kinds of power transmission mechanisms 104, 105 are input to any two of three elements (a sun gear, a carrier, and a ring gear) of the planetary gear mechanism 103, and the remaining one element is connected to an output shaft 106 of the continuously variable transmission apparatus, whereby a differential component between the two elements (for example, the ring gear and carrier) can be output to the output shaft 106.
In a low speed running mode, all of the power (torque) of an engine 107 is transmitted through a drive shaft 108, toroidal-type continuously variable transmission 102 and first power transmission mechanism 104 to the output shaft 106. On the other hand, in a high speed running mode, the power of the engine 107 is transmitted through the second power transmission mechanism 105 and planetary gear mechanism 103 to the output shaft 106, while part of the power is input from the planetary gear mechanism 103 to the output side disk of the toroidal-type continuously variable transmission 102.
The above arrangement can reduce the torque to be applied to the toroidal-type continuously variable transmission 102 in a vehicle high speed running operation, can enhance the durability of the respective parts that constitute the toroidal-type continuously variable transmission 102, and can enhance the torque transmission efficiency of the whole of the continuously variable transmission apparatus.
By the way, in JP-A-11-108147, there is proposed a technique in which the numbers of rotations of the two elements of the planetary gear mechanism are measured and, when the numbers of rotations are almost coincident with each other, the switching operation between the high speed and low speed modes is executed. Also, in JP-A-9-89072, there is disclosed a continuously variable transmission apparatus of a so called geared neutral type in which a toroidal-type continuously variable transmission is combined with a single planetary gear mechanisms. In the continuously variable transmission apparatus of a geared neutral type, in a low speed running operation, power is transmitted through the single planetary gear mechanism and toroidal-type continuously variable transmission; and, in a high speed running operation, the power is transmitted only through the toroidal-type continuously variable transmission. According to the continuously variable transmission apparatus of a geared neutral type, there can be obtained an advantage that, in case where, in a low-speed-side speed mode, the continuously variable transmission apparatus is controlled such that a differential component of the planetary gear mechanism provides zero rotation, there can be eliminated the need for provision of a start clutch.
However, in this type of continuously variable transmission apparatus comprising the toroidal-type continuously variable transmission and planetary gear mechanism combined together, in the mode switching operation, torque to be input to the toroidal-type continuously variable transmission varies greatly from positive to negative (or vice versa). For example, in the switching operation for switching the low speed mode over to the high speed mode, there is a possibility that the input torque can vary from +300 Nm to −240 Nm.
Now, FIG. 19 shows part of a power roller 4 and a trunnion 7a used in the toroidal-type continuously variable transmission 102. In case where a load is applied to the toroidal-type continuously variable transmission 102, there occurs a traction force in the axis X1 direction of the trunnion 7a. In case where the power roller 4 is shifted, for example, in a direction shown by an arrow mark M1 due to the traction force, a feedback mechanism operates in such a manner that it returns the power roller 4 in the opposite direction (a direction shown by an arrow mark M2).
In a radial needle roller bearing 124 for supporting the power roller 4 and in a radial needle roller bearing 125 for supporting a shift shaft 55, inevitably, there exist gaps. Therefore, in case where a load is applied to the toroidal-type continuously variable transmission 102 due to the above-mentioned traction force, the power roller 4 is moved in the axis X1 direction by the sum total amount of these gaps.
Since the power roller 4 is moved in the axis X1 direction for the above reason, the load is applied while the transmission ratio of the toroidal-type continuously variable transmission 102 is left fixed, that is, without issuing a transmission instruction. In other words, in case where the input torque is caused to vary from 0 Nm, as shown in FIG. 20, as the input torque varies, the transmission ratio of the toroidal-type continuously variable transmission 102 is caused to vary, in spite of the fact that no transmission instruction has been issued. That is, the movement of the power roller 4 in the axis X1 direction causes side slippage between the power roller 4 and disk, which causes the power roller 4 to swing and rotate, thereby changing the transmission ratio of the toroidal-type continuously variable transmission 102.
In FIG. 20, the transmission ratio of the toroidal-type continuously variable transmission 102 varies greatly in the low torque area. The reason for this is that, in case where the power roller 4 is moved by the above-mentioned sum total of the gaps in the low torque area, the power roller 4 is caused to swing and rotate. FIG. 20 shows the results obtained from a test conducted under the conditions that the transmission ratio was set about 0.5, the number of rotations was set constant for about 2000, and the temperature of oil was set near to the actual temperature of a vehicle.
Now, description will be given below further of the phenomenon in which the transmission ratio varies mainly in the low torque area in the above-mentioned manner with references to FIGS. 21A, 21B and 22.
FIG. 21A shows a state of a trunnion in which a load applied to the trunnion is zero. In this case, a pair of power rollers respectively supported on their associated shift shafts 55 are both situated at their initial positions (that is, neutral positions); and, trunnions 7a, 7b, following the power rollers, are also situated at their respective initial positions which are (right and left) symmetric. Thus, the precess cam 74 of the feedback mechanism is also held at its given initial position.
FIGS. 21B shows a state of a trunnion in a light load area in which input of the torque is started. In the light load area, since the traction force acts, the power rollers and trunnions 7a, 7b are moved in the axial direction X1 by an amount corresponding to the above-mentioned gaps. Also, because the shift shafts 55 supporting the power rollers are respectively held like cantilevered beams, the flexing amounts of the shift shafts 55 are also included in the moving amounts of the trunnions 7a, 7b. 
In case where the power rollers are moved in the axis X1 direction in the above manner, there is caused side slippage between the disk and power rollers and, in response to this, the feedback mechanism operates, so that the power rollers are finally returned to their respective initial positions (neutral positions). During this operation, since the transmission control valve has not received any transmission instruction, the moving amount of the transmission control valve depends only on the load and thus the transmission control valve little moves. Therefore, the power rollers are swingly rotated by the amount corresponding to the moving amount of the precess cam in the axis X1 direction, and the swung rotational movements of the power rollers cause the transmission ratio to vary.
FIG. 22 shows a state of a high load area. In the high load area, since the shift shafts 55 are flexed, the trunnions 7a, 7b are moved further in the axis X1 direction. However, in the toroidal-type continuously variable transmission of a traction drive type, since the traction force is generated by pressing the disk against the power rollers, due to such pressing force, the trunnions 7a, 7b are elastically deformed in such a manner as shown in FIG. 22 in an exaggerated manner.
With the elastic deformation of the trunnions 7a, 7b, the support portions J1, J2 of each of the trunnions 7a, 7b move in the direction (axis X1 direction) where they approach each other. Because the moving direction of the support portion J2 is opposite to the moving direction of the trunnion 7a caused by the existence of the above-mentioned gap, the moving amounts of these two elements cancel each other. As a result of this, the moving amount of the precess cam 74 in the axis X1 direction is slight. For these reasons, in the high load area, there hardly occurs such swung rotation of the power roller (that is, the variation of the transmission ratio) as occurring in the low load area.
Also, there can also be found a phenomenon in which, with a load applied to the toroidal-type continuously variable transmission, the trunnions 7a, 7b are elastically deformed and trunnion shafts 119 are flexed, so that the transmission ratio of the toroidal-type continuously variable transmission. In the toroidal-type continuously variable transmission of a traction drive type, it is necessary to press the disk against the power rollers and the pressing forces are supported by the trunnions 7a, 7b. The trunnions 7a, 7b, as one trunnion 7a is representatively shown in FIG. 19, are supported by a pair of support members 49, 49 which are referred to as yokes, whereby mutually-opposite-direction forces generated between the respective trunnions 7a are allowed to cancel each other. Due to this, each of the trunnions 7a is elastically deformed between the pair of support members 49, 49.
The trunnion shaft 119 itself is not deformed elastically because it does not receive the above pressing force; however, due to the influence of the above-mentioned elastic deformation caused between the support members 49, 49, the trunnion shaft 119 and rod 73 are swung. Due to this, the contact point between the precess cam 74 and the cam follower of the transmission control valve is caused to vary, so that the spool of the transmission control valve is moved in the axial direction.
As a result of this, the transmission control valve is operated to generate the oil pressure that moves the trunnions 7a, 7b in the axis X1 direction, thereby changing the transmission ratio of the toroidal-type continuously variable transmission. For example, in case where the precess cam 74 is moved in the axis X1 direction, the power roller is swingly rotated by an amount corresponding to the moving amount of the precess cam 74 according to the cam lead. For instance, in the case of the cam lead being 20 mm/360°, in case where the precess cam 74 moves by 0.3 mm in the axial direction X1, the power roller is swingly rotated no less than 5.4°.
In case where these factors are combined together, as shown in FIG. 23, although no transmission instruction is issued, as the torque varies, the transmission ratio of the toroidal-type continuously variable transmission is caused to vary. Accordingly, even in case where, as disclosed in the above-cited publication JP-A-11-108147, the clutch is connected when the number of rotations of the toroidal-type continuously variable transmission is coincident with the number of rotations of the planetary gear mechanism, as the torque varies, the transmission ratio of the toroidal-type continuously variable transmission is caused to vary greatly, thereby causing the number of rotations of the engine to vary.
By the way, in case where the driver judges the need for engine braking and takes his or her foot off the accelerator, similarly to the above case, although no transmission instruction is issued, the torque is caused to vary suddenly In this case as well, the transmission ratio of the toroidal-type continuously variable transmission is caused to vary. However, in this case, since the engine braking is used of the driver's will, the driver can forgive a certain degree of transmission shock caused by such variations in the transmission ratio of the toroidal-type continuously variable transmission.
However, in the continuously variable transmission apparatus of the above-mentioned power circulation type, even in case where the driver does not intend to switch the transmission mode, the mode is switched involuntarily. Therefore, in case where even a slight degree of transmission shock occurs in the mode switching operation, the drive feels incongruous. That is, as in the conventional the toroidal-type continuously variable transmission (FIG. 23), in case where, when no transmission instruction is issued in the mode switching operation, the transmission ratio varies greatly to thereby cause a transmission shock, the driver feels badly incongruous.